GB2136055A - Hydraulic Feed Circuits for Actuators - Google Patents

Abstract

Hydraulic circuits are disclosed which connect a pump 13 to an actuator 11. The pump delivery is maintained at low pressure by a regulating valve 28 when the actuator 11 is at rest, in order to limit the dissipated power. When a control valve 21 is operated by a lever 23 to supply the actuator 11, a pilot circuit 19 is cut off from exhaust and is pressurised by the pump to load the valve 28 and increase the pressure in the pump delivery line 15. The pilot circuit 19 is pressurised by the pump 13 through a constriction 20. The line 19 is connected to a pilot line 31 containing a non- return valve 33 and constriction 32, the line 31 being branched from the line 24 to the actuator downstream of the valve 21. The restrictions 20, 32 reduce the pressure difference between lines 16 and 24 thereby avoiding sudden movements and pressure changes. <IMAGE>

Description

SPECIFICATION Hydraulic Feed Circuits for Actuators Actuator feed circuits are known which are fed by a pump having a continuous delivery which is fed to the actuator when the actuator is to be driven, whereas it is fed to discharge when the actuator is to remain inoperative.

One problem of these circuits is to limit the pump delivery pressure when the delivery is connected to discharge. This can be done by connecting the pump delivery to discharge by way of a pressure relief valve which is made to close by a pilot signal taken from the actuator feed pipe.

This leads to drawbacks because pilot liquid has to be withdrawn in order to close the pressure relief valve to increase the delivery pressure, and this can cause an initial lowering of the load which is being raised.

In addition in spite of relative complexity of the control members for both the working and pilot flow, it is not possible to control the throughputs and pressures with the required accuracy and progression.

These known circuit arrangements are also of little flexibility, in the sense that their components and calibration must to a large extent be modified in order to adapt to the various circuit powers.

The object of the invention is to reduce the reserve or loss of inoperative pressure which occurs in the aforesaid hydraulic systems but without leading to a reduction in the operating drive pressure which could otherwise prejudice the stability and quality of the drive.

A further object of the invention is to obviate two important drawbacks of compensated controlled force pressure circuits in a simple and effective manner, these drawbacks being: a) the need to determine with high accuracy, normally by using costly spool valves, the time of initial movement of the pressure compensating signal or load sensor, and the need for control means which direct or limit the flow generated by the variable volume pump or compensator pump control means; b) the need to actually withdraw a pressure signal from the load circuit, which in practical terms means withdrawing a volume from the circuit which supports the load, this causing a small but unacceptable initial movement in the opposite direction to the required movement.

A further object of the invention is to allow a reduction in both the constructional and operating costs of classical follow-up servo systems, such as could easily be used in a control system for an agricultural tractor implement.

A further object is to enable the increase and reduction in the flows fed to the load to be controlled on a time basis by hydraulic control, and to enable low-cost valve means to be used even in high inertia systems, without undue pressure peaks being produced, as are normally associated with this type of control.

These objects are attained according to the invention by a hydraulic control circuit for a hydraulic actuator, comprising a pump which feeds liquid under pressure to the actuator through a pipe provided with means for regulating the quantity of liquid fed to the actuator, and which is also connected to discharge through a branch incorporating a pressure regulating valve which is made to open by a signal proportional to the pump delivery pressure, in opposition to a fixed preload and also proportional to the action of the pressure in an auxiliary pilot circuit which withdraws-liquid from the delivery pipe by way of a sized port and which can be connected either to discharge or to the actuator inlet pipe by way of a non-return valve.

The objects and characteristics of the invention together with its practical application will be apparent from the description given hereinafter of circuits which use the principles of the invention, and which are illustrated on the accompanying drawings in which: Figure 1 is a diagram of a control circuit using a slide valve; Figure 2 is a diagram of a circuit which is controlled by acting on the pilot flows; Figure 3 shows a component of the diagram of Figure 2 in greater detail; Figure 4 is a modification of the component of Figure 3; Figure 5 is a modification of the circuit shown in Figure 2.

With reference to Figure 1, a hydraulic control circuit for a hydraulic actuator 11, for example a cylinder moving a load 12, comprises a pump 13 connected to a sump 14 and to a main delivery pipe 1 5, and a pressure relief valve 1 6.

The main pipe 1 5 is connected to a pipe 17 provided with a unidirectional valve 18, and to a pilot line 1 9 provided with a first constriction 20.

A three-position slide valve 21 with a flow regulating section 22 is operated by a manual control lever 23.

Said valve 21 controls the pipe 17, a pipe 24 connected to the actuator 11 , a pilot line 27 connected to the pilot line 19, and the pipe 25 and line 26 leading to discharge.

A second pressure regulating valve 28 is kept closed by an elastic means 29 and by the pressure of the pilot line 19.

A pilot line 31 connects the pipe 24 to the lines 1 9 and 27, and is provided with a second constriction 32 and a non-return valve 33.

The circuit of Figure 1 is shown in its rest position with a fixed volume pump 13 feeding fluid to the pipe 15 and then to the line 1 9 to pass through the first constriction 20.

The pressure regulating valve 28 is acted upon by the low-load spring 29 and also by the pressure of the line 19, which can be assumed to be zero in this position as this line is connected to discharge by way of the lines 26 and 27.

Under these conditions, the pressure regulating valve 28 is sensitive to the pressure in the pipe 15, to open to discharge when this pressure reaches the value defined by the loading of the spring 29, which is low in order to limit the power dissipated when at rest.

If the lever 23 is now moved towards the left, the line 27 closes, and the fluid fed by the pump 1 3 into the line 1 9 generates an increasing pressure which tends to close the valve 28 in cooperation with the loading of the spring 29. The flow begins to pass through the pipe 1 7 and through the valve 21 to the hydraulic actuator 11 to cause it to rise. The speed and acceleration of its rise are regulated by the loading of the spring of the non-return valve 33 and of the spring 29, and by the constrictions 20 and 32 which positively change the AP between the pipe 24 and pipe 17, so reducing it in order to prevent sudden movements and pressure changes.

On moving the lever 23 towards the right into its neutral position, the pipe 24 is connected to the pipe 25 and to the sump 14, to cause the hydraulic actuator 11 to descend, the lowering movement being regulated by the variable orifices 22. The line 27 is connected to discharge, with consequent reopening of the valve 28.

Further advantageous economical simplifications can be obtained by replacing the three-position slide valve 21 in the manner shown in Figure 2.

This circuit is equivaient to that of Figure 1 from many aspects, and those components corresponding to the components of this latter circuit are indicated by the same reference numerals. In the circuit of Figure 2 the unidirectional valve 1 8 is replaced by a non-return valve 34 provided with elastic means which determine a load exceeding that determined by the spring 29 on the valve 28.

The line 27 incorporates a first valve element 35 operated by a lever 36 which also regulates the pilot flows by way of a second valve element 37.

The unit for discharging liquid from the actuator is indicated overall by 46.

A valve 38, normally closed by a spring 39, connects the pipe 24 to the pipe 25 leading to discharge, and is opened by a stem 40 rigid with the piston 41 of a double-acting cylinder 42.

Said piston 41 moves in one direction or the other under the thrust of a pressure proportional to that of the actuator 11 and present in a line 43, and of a pressure regulated by the valve element 37 and present in a line 44, a connection provided with a constriction 45 being provided between the two lines 43 and 44.

Replacing the three-position slide valve 21 with its variable orifices 22 by said simple valve elements and valves enables the same functions to be performed as described heretofore.

In this respect, in'the position shown in Figure 2 with the lever 36 in its middle position, the fluid in the line 27 is connected to discharge and induces a pressure approximating to zero in the line 19, to enable the pressure regulating valve 28 to discharge the pump flow.

On moving the lever 36 towards the right so as to completely close the valve element 35, the pressure is increased in the lines 19, 27 until the pressure regulating valve 28 is closed, thus causing the hydraulic actuator 11 to rise, the rise being regulated by the means provided in the manner heretofore described.

On moving the lever 36 into its left hand end position, both the valve element 35 and the valve element 37 are opened to discharge in order to cause the actuator 1 to descend.

The return flow from the pipe 24 passes into the pilot line 43 to drive the piston 41 and open the valve 38 to discharge through the pipe 25.

On closing the valve element 37 again by means of the control lever 36, said sequence is reversed and the valve element 38 is closed.

Figure 3 shows an improvement relative to the block 46 of Figure 2. A hollow piston 48 of crosssectional area A1 slides in a cylinder 47 against a spring 49, in order to close a seat 50 with one of its ends, which is of area A2 ( < A1).

The wall of the hollow piston 48 comprises a sized port 51 which connects a rear chamber 52 to a front chamber 53. The first chamber 52 is connected to the line 44 the discharge of which is regulated by the valve 37, and the second chamber 53 is connected to the feed pipe 24 of the actuator 11.

The piston 48 extends in the form of a stem 54 housed in a third chamber 55 connected to discharge and carrying a piston 56 which defines a chamber 57 fed by a pipe 58 with fluid at a pilot pressure Pp The valve 37 is made to provide two levels of pilot signal by providing it with a head 59 which is operated by the lever 36 and is of slightly smaller diameter than the valve bore 60, in order to provide two discharge positions, in the first position the valve 37 being open and the head 59 remaining inside the bore in order to form a constriction, whereas in the second position the head 59 has completely emerged from the bore to give free discharge.

A more simple embodiment of the scheme of Figure 3 can comprise a valve 37 of the type shown in Figure 2, and the pipe 58 can be simply connected to discharge downstream of the constriction 47, thus operating with Pp equal to zero.

The operation of the block 46 is more apparent from the following considerations: Ignoring the friction forces and the effect of the flow forces which can both be reduced to small values by methods well known to the art, the point of static equilibrium of such a valve 47 can be considered to satisfy the following equation: P0(A1-A2)=P1A1 +spring force so that when there is no flow through the sized port 51, P,=PO and the valve remains closed at all pressures. An increase in flow on the part of P0 through the sized port 51 to discharge by way of the controlled valve 37 reduces P, to a value corresponding to the opening of the valve.

Providing the extension 56 of the piston 48 in the valve represents one way of overcoming the equilibrium force variation which occurs when the area A2 senses a downstream pressure variation in the discharge line, whilst the constriction 67 provides a convenient way of reducing the closure speed of the valve 47, if necessary, by closing the valve 37 when the flow through the orifice 51 has stopped.

It remains only to influence the flow control in a more reasonable manner than that obtainable with a pilot valve with its inherent stability, this normally being done by applying a high elastic force or creating a small surface area difference (A1-A2), in order that a small variation in the pilHlflow which causes a small change in PiP, does not excessively vary the closure of the valve in its seat 50. These high differential forces or elastic forces would limit the minimum pressure at which the valve can open, but these performance limitations can be overcome in the manner described hereinafter, the utility of the special means used being clear with reference to Figure 3.

On applying a constant pilot pressure P to the left hand side of the piston 56, a greater elastic force can be applied in order to close the valve thus requiring a greater PO/Pa ratio, and enabling the valve gain to be controlled at will.

It has been found that in practice that a relatively rigid spring 49 assembled with a small preload enables a modest control pressure to balance the opening and closure forces in a position in which the valve element is just open.

Under such conditions any depression can be made up in the actuator loading circuit when the pilot valve 37 is open.

Closure of the pilot valve 37 balances P0 and P,, and the valve closes under the influence of a very small loading pressure PO.

It has been found advantageous in the invention to provide two levels of pilot signal or downward movement control, these being obtained by the head 59.

When the lever 36 is moved into the intermediate lowered position, by which the pilot valve means 37 is commanded to carry out the descent, the movement enables the valve ball to rise from its seat but the head 59 of the operating rod remains in the pilot valve bore 60, in such a manner as to limit the flow from the actuator 11 for the load 12 which passes through the constriction 51 in the valve 47, so that P, does not fall sufficiently to open the main valve. The small flow passes through the line 44 and line 26 to the sump.

This small flow is suitable for finely controlling the position of the actuator 11 and for attenuating any acceleration or deceleration surges in the loading system.

When the pilot valve 37 is moved by the lever 36 into its completely lowered position, the head 59 of the operating rod emerges from the bore 60, and the flow to the sump from the loading circuit which now passes through the orifice 51 in the valve 47 is sufficient to lower P, and open the valve.

Designing the operating rod head 59 in accordance with the current state of the art allows flow conditions which are substantially turbulent or insensitive to viscosity at this point, and also allows flow modulation to be obtained in order to provide smooth closure or opening of the main valve.

in many applications, a control means will be required for limiting the maximum descent or discharge rate independently of the value of the driven load or of the means for signalling the descent.

The principle well known to the art uses a compensated pressure flow control valve connected in series with the load so that substantially the entire flow from the loading circuit has to pass through said valve art a controlled maximum throughput.

The use of such a valve has been considered as an optional characteristic of the invention and would be used with the type of descent control valve 47 which remains substantially uninfluenced by the use of a flow control valve downstream, i.e. between said descent valve 47 and the sump. However a further novelty which is valuable in increasing the versatility and reducing costs is to combine the function of the controlled descent valve with a flow control valve so that the controlled descent valve can be immediately replaced by the combined valve without modifying the control means.

Figure 4 shows this novelty, which operates in the following manner: The already described descent valve 47 is given a further passage path 61 which connects the pressure acting in the pipe 62, between the seat 50 and an orifice 63, to that chamber 52 of the valve 47 of cross-section A,.

A constriction 64 limits the flow from the pipe 62 to A1, and the passage 61 is closed in order to discharge a flow passing from the pressurised loading circuit through the constriction 51 by way of a ball valve 65 kept in position by a spring 66.

When the pilot valve is opened, descent occurs in the following manner: As stated heretofore a small flow can pass through the valve 37 without opening the main valve 47, but when the pilot means 37 is completely open, the valve 47 opens in the manner heretofore described.

If the flow from the loading circuit to the sump by way of the valve 47 and the orifice 63 does not reach a predetermined throughput, the pressure difference across the orifice 63, i.e. P2-P3, does not enable P2 to exceed the pressure P, in the chamber A1, and the non-return valve 65 remains closed.

Above a certain pressure PO, i.e. above a certain loading on the actuator 11 for the load 12, the value of P2 increases beyond that of P1, so that the non-return valve 65 opens, and the increased flow to the pilot valve gives rise to an increase in the pressure P, which begins to close the valve 47 against the seat 50 until P1, the effect of the spring 66, and the orifice 28 all combine to balance the effect of P2.

The invention does not exclude means for varying the value of the orifice 63 under the control of the system operator, so that a limitative speed range can be used according to the particular case. The valve combination can also be used for a different method of providing continuous modulation if this can be advantageous, in the following manner.

By eliminating th-e non-return valve 65 and the elastic means 66, closing the line to the pilot valve 37 and eliminating the valve 37 itself, the orifice 63 can be made variable by known means, to control the closure pressure P, of the valve 47 by the flow from the loading circuit P1 by way of the orifices 51 and 60.

In this manner each extent of opening of the orifice 63 maintains the equilibrium AP at the valve P2-P0, and the descent rate remains substantially constant independently of the load.

The aforesaid improvements can be used with advantage in servo systems for tractor implements or lifting systems where it is very important to attain low costs, increased reliability, flexibility and efficiency in use.

In response to one or more combined signals from various sources, a tractor lifting system responds by lifting or lowering a mass by means of a hydraulic actuator, generally a single-acting piston, and the control circuit would therefore be included in the working flow system.

This flow is usually generated by fixed volume pumping means, so that that part of the flow which is not actually used for moving the load is fed to the sump, so wasting an amount of energy proportional to the pressure required in the drive means for lifting the load, and which in a system in which the pumped pressure is relative and limited by the load pressure is proportional to the time for which the pumping means is under pressure and to the force lost.

Figure 5 shows a servo system applied to an agricultural tractor. In this figure the input to the system is mechanical, the movement being converted into a hydraulic pressure signal by means of a spool valve 68 and the orifices 69 and 70 in the feed lines from a low-pressure source P but it can also relate to other combinations which provide the same input signals.

In this embodiment, the functions of the controlled valves 35 and 37 of the preceding figures are performed by units 75 and 84 respectively, which are servo-controlled by a very low pressure circuit derived by way of the pressure reducing valve 95.

When in the non-operating position the flow from the low pressure source passes through the orifices 69, 70 to the sump by way of the large passages 71 and 72 in the body of the spool valve 68. When the valve 68 is moved by a lever 73 towards a left hand position A, the pressure rises in a chamber 74 of a cylinder 75 by the effect of the throttling or closure of the passage 71, and a piston 77 commences movement against the action of a spring 76.

The piston 77 has a sized orifice 78 which passes through it and emerges into a valve seat 79. The flow enters the chamber 74 and passes through the orifice 78, and the flow from the valve seat 79 to the sump limits the pressure in the chamber 74 to that generated by the reaction of the flow originating from the line 27 and acting by way of a pin 80, because the seats 79 and 81 are equal.

It does not matter if the pilot valve 68 causes the pressure in the chamber 74 to rise more or less slowly, because at a certain point the flow passing through the orifice 78 ceases to cause the pin 80 to close the seat 79, and the pressure in the chamber 74 rises immediately, without possibility of modulation, to reach the maximum pilot pressure which produces a force on the piston 77 sufficient to cause the pin 80 to close the seat 81, to begin to increase the positive return pressure which loads the valve 28 as described heretofore.

On returning the lever 73 to its non-operational neutral position, the valve 68 frees the passage 71 and begins to pass flow to the sump, to reduce the pressure in the chamber 74 until the seat 81 is not open and the pressure signal controlled by the load applied to the valve 28 falls to a value below that of the loading pressure, and the flow ceases.

The pressure in the chamber 74 falls progressively to zero, to ensure that the flow passing through the seat 81 and the pin 80 is also essentially at zero pressure.

In controlling the descent movement (lever 73 in position B), the pilot valve 68 partially closes the passage path 72 towards the sump, to give rise to an intermediate pressure less than the pilot pressure which has to be generated proportional to the areas of the constriction 70 and a constriction 82 provided in the body of the spool of the valve 68.

This intermediate pressure when applied to a chamber 83 of a cylinder 84 is sufficient to force a piston 85 and a stem 86 against a pin 87 which in its turn pushes against the ball of the valve element 37. The seat 88 for the ball of the valve element 37 has a greater diameter than the pin 87.

In this manner when the thrust on the piston 85 exceeds the force of the load which acts on the diameter of the seat 88, the effect of the loading pressure immediately reduces by acting only on the lower area of the pin 87, thus ensuring that no modulation relative to the input in either the opening or closure mode is possible.

When it is opened in this manner, the flow which passes from the actuator 11 for the load 12 through the orifice 51 and into the valve 47 is limited to below that necessary for lowering the pressure to a sufficient extent for opening valve 47, because of the limiting effect of the head 59 of the pin 87 which remains in the constriction 60.

In this respect, a preloaded spring 89 which operates against a disc 90 urged against a stop 91 is sufficient to oppose the further movement of the ball of the valve element 37 in the opening direction.

When the spool of the pilot valve 68 is further moved (position C of the lever 73) in order to completely close the passage 72 towards discharge, the maximum pilot pressure acts on the piston 85. The pressure increase overcomes the resistance of the preloaded spring 89, and the stem 86 slides inside the disc 90, which is locked by abutting against the stop 91.

The stem 86 urges the pin 87 further forwards until the head 59 completely emerges from the constriction 60 to allow a considerable flow through the valve element 37.

The flow of the actuator 11 for the load 12 now increases through the orifice 51 as heretofore explained, so reducing P, and enabling the main valve 47 to open.

The operational sequence is reversed by the pilot piston 68 moving in the opposite direction.

The embodiments illustrated in the schemes described heretofore are given as examples only, and the principles on which the present invention is based can find application in different circuits, for adaptation to specific uses.

Claims (10)

1. A hydraulic control circuit for a hydraulic actuator, comprising a pump which feeds liquid under pressure to the actuator through a pipe provided with means for regulating the liquid flow to the actuator, and which is also connected to discharge through a branch incorporating a pressure regulating valve, characterised in that said pressure regulating valve is made to open by a signal proportional to the pump delivery pressure, in opposition to a fixed preload and also proportional to the action of the pressure in an auxiliary pilot circuit which withdraws liquid from the delivery pipe by way of a sized port and which can be connected either to discharge or to the actuator inlet pipe by way of a non-return valve.

2. A circuit as claimed in claim 1, characterised in that said non-return valve in said auxiliary circuit opens against a predetermined load and is also associated with a constriction.

3. A circuit as claimed in claim 1, characterised in that said means for regulating the quantity of liquid to the actuator are constituted by a valve comprising a valving member and a flow regulating section.

4. A circuit as claimed in claim 1, characterised in that said means for regulating the liquid flow to the actuator are constituted by a first valve element which controls the connection of said auxiliary pilot circuit to discharge or to the actuator feed pipe, and by a second valve element which controls the connection of the actuator inlet pipe to discharge.

5. A circuit as claimed in claim 4, characterised in that said second valve element comprises a valve urged into its closed state by a preload and controlled by a second auxiliary pilot circuit comprising a first line connected to the actuator inlet pipe to provide a pressure signal to cause opening against the action of the pressure in a line which can be connected either to discharge or, by way of a constriction, to the actuator inlet pipe.

6. A hydraulic circuit as claimed in claim 5, characterised in that said second valve element is constituted by a differential valve in which a member of larger diameter slides in a cylinder urged by a spring in order to cause a valving member of smaller diameter to rest against the corresponding seat, said member of larger diameter defining a rear chamber and a front chamber in which the valving member moves and which is connected to the actuator feed pipe, the front and rear chambers being connected together by a sized port, the rear chamber being openable to discharge by said means for regulating the quantity of liquid fed to the actuator.

7. A circuit as claimed in claim 6, characterised in that said cylinder is mechanically connected to a piston member mobile in a seat in order to define a further chamber which is fed by liquid through a line provided with a constriction.

8. A circuit as claimed in claim 6, characterised in that said further chamber is fed with pressurised liquid to determine a thrust on the piston acting in said cylinder in opposition to said elastic means.

9. A hydraulic circuit as claimed in claim 6, characterised in that said valving member controls the discharge by way of a constriction, and the discharge between the valving member and the constriction is connected by way of a non-return valve to said rear chamber.

10. A hydraulic circuit as claimed in claim 6, characterised in that said rear chamber can be opened to discharge by way of a valve, the degree of opening of which can be varied manualiy.